energies

Article A Design of the Compression Chamber and Optimization of the Sealing of a Novel Rotary Internal Combustion Using CFD

Savvas Savvakis * , Dimitrios Mertzis , Elias Nassiopoulos and Zissis Samaras

Department of Mechanical Engineering, Aristotle University of Thessaloniki, 541 24 Thessaloniki, Greece; [email protected] (D.M.); [email protected] (E.N.); [email protected] (Z.S.) * Correspondence: [email protected]

 Received: 29 March 2020; Accepted: 1 May 2020; Published: 9 May 2020 

Abstract: The current paper investigates two particular features of a novel rotary split engine. This internal combustion engine incorporates a number of positive advantages in comparison to conventional reciprocating piston . As a split engine, it is characterized by a significant difference between the expansion and compression ratios, the former being higher. The processes are decoupled and take place simultaneously, in different chambers and on the different sides of the rotating pistons. Initially, a brief description of the engine’s structure and operating principle is provided. Next, the configuration of the compression chamber and the sealing system are examined. The numerical study is conducted using CFD simulation models, with the relevant assumptions and boundary conditions. Two parameters of the compression chamber were studied, the intake port design (initial and optimized) and the sealing system size (short and long). The best option was found to be the combination of the optimized intake port design with the short seal, in order to keep the compression chamber as close as possible to the engine shaft. A more detailed study of the sealing system included different labyrinth geometries. It was found that the stepped labyrinth achieves the highest sealing efficiency.

Keywords: SARM; rotary engine; compression chamber; combustion chamber; sealing; CFD

1. Introduction

Aiming at the mitigation of the Greenhouse Effect, stringent CO2 emission regulations and targets have been set, for passenger cars and light commercial vehicles (e.g., in the European Commission [1] and in the US [2]), as well as heavy-duty vehicles [3]. Such targets pose strong challenges to the automotive sector. Various pathways can be followed in order to address these challenges. Although electrification is probably the most widely applied solution, it is only the battery electric vehicle (BEV, charged by the grid, or FCEV) that is completely independent from an “onboard” thermal engine and becomes completely independent from thermal energy only in the case that grid electricity (or hydrogen for FCEV) is produced by renewable sources (sun, wind, water, etc.). In addition, there is a number of key factors (such as battery cost, recharging infrastructure network, and CO2 car standards) that greatly influence BEV’s market penetration [4]. On the other hand, hybrid vehicles (HEV, PHEV, range extender) are still equipped with a thermal engine. Another pathway is fuel decarbonisation. In this case, alternative fuels with lower carbon content can be used (e.g., natural gas, consisting mainly of methane), while also (green) e-fuels are developed, which are synthetic fuels produced by the extensive use of renewable energy, thus reducing CO2 emissions of the complete fuel life cycle [5]. Independently, however, of the fuel production process

Energies 2020, 13, 2362; doi:10.3390/en13092362 www.mdpi.com/journal/energies Energies 2020, 13, 2362 2 of 21 and properties, this is used in a thermal engine, in a variety of applications ranging from road transport to shipping [6]. It is clear from the above that the internal combustion engine (ICE) still plays, and will continue to play, an important role in transportation systems. Thus, it is of great importance to make the internal combustion engine as efficient as possible, always keeping the right balance between fuel economy and pollutant emissions. To that aim, various roadmaps have been developed around the globe towards the development of highly efficient thermal engines, with the efficiency target reaching 50% in passenger car applications [7] and 55% in heavy-duty applications (US Super Truck program [8,9]), while currently the peak performance in passenger cars is in the order of 40% [7]. As a result, considerable resources are invested in developing fuel economy and boosting technologies, optimising engine combustion strategies and thermal management, using alternative operating cycles like the Atkinson, inventing new concepts and technologies, developing controllers and others [10]. However, most engines operate at less than half of the theoretical efficiency limit [11]. The main development activities focus on the reciprocating piston ICE, which, however, presents a number of disadvantages in terms of engine dynamics, due to the necessary conversion of the reciprocating to rotational motion. The latter results in bigger and heavier constructions, suffering from oscillations and producing only part of the theoretically available power and torque. On the other hand, there is no intense research on rotary engines, resulting in only one commercially available engine for passenger cars, the (four stroke) Wankel engine, also used in other applications (e.g., aviation, motorsport) [12], while its two-stroke version is also examined in combination with supercharging aiming at higher power density [13]. The main drawbacks of this engine are the combustion chamber shape (high surface to ratio at ignition) and the leakage of the working medium between the adjacent chambers (blow-by effect) [14]. In order to overcome the disadvantages of the existing ICEs, various alternative concepts have been developed, such as the Revetec engine [15], the Scuderi Split Cycle (SSC) [16], the Duke engine [17] and others. An overview of such alternative concepts can be found in [18]. In the current , a novel rotary ICE is presented, describing its novel design and operating principle, targeting higher thermal efficiency. The configuration of this engine, called the Savvakis Rotary Motor (SARM), aims at eliminating the drawbacks of the existing ICEs. Its operation is based on the , allowing for improved performance. The engine is concentric and the processes are realized in separate chambers with individual pistons. The objective of the present work is to investigate in more details two particular features of this novel engine, namely the compression chamber design and the sealing system. To that aim, mathematical models are setup in commercial CFD software (Computational Fluid Dynamics), commonly used in such investigations [19], combined with the relevant assumptions and boundary conditions. After the description of the structure of the engine and its operating principle, the computational tools are presented, including the meshes and the parameters examined. The results section highlights the advantages of the different configurations, concluding with the optimum solution.

2. Materials and Methods

2.1. Engine Description This section provides a brief description of the technical features and its operation. The detailed technical description of the engine and its operating principle, together with the justification of the selection of geometrical features, is shown in [20]. Figure1 presents the layout of the engine, where its main technical characteristics are shown. The advantages of the SARM engine are presented in Section 2.1.4. Energies 2020, 13, 2362 3 of 21

2.1.1. Engine Components With reference to Figure 1: • The engine consists of two concentric toroidal rings of significantly different diameter. At the common center of the two rings the engine shaft (1) (power output) is located. • The inner ring (2) forms the intake (2α) and compression (2β) chamber (CPC = ComPressionChamber), while the outer one is the combustion (CBC = ComBustionChamber) (4) and exhaust (5) chamber. • The CPC and the CBC communicate through an intermediate chamber, which is called the Chamber (PC) (6). • Gas exchange between the chambers (from the CPC to the PC and then to CBC) is achieved with the use of sliding ports and valves. • Within each chamber, a pair of symmetrically located pistons rotates; two pistons in the CPC (7) and two pistons in the CBC (8). • A rotating moving arm (or disc) (9) interconnects the CBC pistons, the CPC pistons and the shaft of the engine. All the moving parts of the engine perform rotating motion. • Two additional major components of the engine are the fuel injector and the spark plug. The former can be placed either in the PC or the CBC, depending on the fuel injection strategy, while the latter is placed exclusively in the CBC. • Additional peripheral components include the fuel supply lines, the electric/electronic components for operation and control, engine mounts, etc. Although such components are Energies 2020necessary, 13, 2362 for the operation and control, they are not related with the basic operating concept3 of 21 the engine. Therefore, they are not described further here.

FigureFigure 1. 1.Savvakis Savvakis Rotary Rotary MotorMotor (SARM)(SARM) engine layout. layout.

2.1.1.2.1.2. Engine Operating Components Concept WithFrom reference a thermodynamic to Figure1: point of view, the operating cycle of the SARM engine is based on the Atkinson cycle, which can theoretically reach up to 20% higher efficiency than the [21]. The engine consists of two concentric toroidal rings of significantly different diameter. At the • The major characteristic of an engine running on the Atkinson cycle is that the expansion ratio is common center of the two rings the engine shaft (1) (power output) is located. higher than the compression ratio. This is already applied in reciprocating The inner ring (2) forms the intake (2α) and compression (2β) chamber (CPC = ComPression • piston engines; Toyota Prius has an effective compression ratio (CR) of 8:1 and an expansion ratio Chamber), while the outer one is the combustion (CBC = ComBustionChamber) (4) and exhaust (5) chamber. The CPC and the CBC communicate through an intermediate chamber, which is called the Pressure • Chamber (PC) (6). Gas exchange between the chambers (from the CPC to the PC and then to CBC) is achieved with • the use of sliding ports and valves. Within each chamber, a pair of symmetrically located pistons rotates; two pistons in the CPC (7) • and two pistons in the CBC (8). A rotating moving arm (or disc) (9) interconnects the CBC pistons, the CPC pistons and the shaft • of the engine. All the moving parts of the engine perform rotating motion. Two additional major components of the engine are the fuel injector and the spark plug. The former • can be placed either in the PC or the CBC, depending on the fuel injection strategy, while the latter is placed exclusively in the CBC. Additional peripheral components include the fuel supply lines, the electric/electronic components • for operation and control, engine mounts, etc. Although such components are necessary for the operation and control, they are not related with the basic operating concept of the engine. Therefore, they are not described further here.

2.1.2. Operating Concept From a thermodynamic point of view, the operating cycle of the SARM engine is based on the Atkinson cycle, which can theoretically reach up to 20% higher efficiency than the Otto cycle [21]. The major characteristic of an engine running on the Atkinson cycle is that the expansion ratio is Energies 2020, 13, 2362 4 of 21 higher than the compression ratio. This thermodynamic cycle is already applied in reciprocating piston engines; Toyota Prius has an effective compression ratio (CR) of 8:1 and an expansion ratio (ER) of about 13:1. After testing, Toyota concluded that the engine running on the Atkinson cycle can be 12%-14% more efficient than the corresponding engine running on the Otto cycle [22]. Compared to conventional reciprocating engines, where all processes take place within the engine cylinder (otherwise called the combustion chamber), but during different strokes, in this engine, all processes occur simultaneously, but in separate chambers. These types of engines are called split engines. On the other hand, all chambers, the piston’s rotation and the shaft are located around the same axis, and from this point of view, it could be said that the SARM engine presents some similarities with gas turbines. Both are concentric internal combustion split engines. The presence of the pressure chamber between the compression and the combustion chambers allows for the independent optimization of each (intake, compression, combustion, expansion). Decoupling the processes and the volumes where they take place offers great flexibility, e.g., on the shape and the material of each chamber, or even on the surface treatment of each chamber. From the dynamics point of view, the SARM engine also presents some particular features. Compared to conventional reciprocating engines, the SARM engine eliminates any reciprocating motion in all directions, thus limiting oscillations and confining vibrations, mechanical losses and additional inertial forces, which increase the thickness and weight of an engine’s component. Furthermore, the concentric operation of the SARM engine avoids the additional inertial forces developed by eccentricity in other rotary engines, permitting also higher piston speeds that increase the power density. Finally, the symmetric location of pistons balances the engine, neutralizing oscillations due to centrifugal and inertial forces, while in conventional reciprocating engines a system of counterweights is usually needed to balance the crankshaft. The generated symmetry of the SARM engine eliminates the mechanical vibrations and diminishes the sound effects.

2.1.3. Phases of Operation The SARM engine phases of operation are depicted in Figure2, but there is also a video that describes them which can be seen in the Supplementary Materials.

The CPC piston (7) rotates with a speed of 100 to 12,000 rpm inside the inner ring (2) which has a • toroidal shape. The piston (7) and the sliding port (SP) (3) divide the inner ring (2) into two chambers; the intake • chamber (2α) and the compression chamber (2β). The back side of the piston (7) drags atmospheric air through the intake port (10) into the intake • chamber (2α), while the front side of the piston compresses the trapped air between the piston and the SP inside the compression chamber (2β). The SP (3) remains closed during the whole compression process. When the piston reaches the • closed SP (3), the latter opens letting the piston to pass and then closes again. Valves (11α and 11β) open 15 degrees before the SP (3) opening and the compressed air is delivered • into the combustion chamber (CBC) (4) through the pressure chamber (PC) (6). The air transfer process is terminating with the closing of valve (11β). The combustion chamber • (4) is now delimited by the closed expansion sliding port, the closed valve (11β), the toroidal shell and the CBC piston (8). Inside the combustion chamber, the processes of fuel injection and mixing with the air take place. • Combustion is initiated by a spark plug initiates the combustion process. The following expansion of the burning mixture exercises a force on the CBC piston (9) producing • work and, through the rotating moving arm (9), transfers the rotating motion and the work to the engine shaft and the CPC piston (7). At the same time, the other side of the CBC piston (8) pushes the burned gases of the previous • combustion cycle to the atmosphere, through the exhaust port (12). Energies 2020, 13, 2362 5 of 21 Energies 2020, 13, 2362 5 of 21

(a) intake and compression process (b) air transfer to the combustion chamber (CBC)

(c) ignition-expansion process

Figure 2. SARM engine operating phases. Figure 2. SARM engine operating phases By this brief description of the SARM engine operation, a number of its special features By this brief description of the SARM engine operation, a number of its special features become become apparent, which form the basis of its advantages compared to the conventional reciprocating apparent, which form the basis of its advantages compared to the conventional reciprocating piston piston engines. engines. A major difference between the SARM engine and all conventional SI engines (either reciprocating A major difference between the SARM engine and all conventional SI engines (either orreciprocating rotary) concerns or therotary) operation concerns of thethe pistons. operation While of the inconventional pistons. While four-stroke in conventional reciprocating four-stroke engines thereciprocating same piston engines performs the periodicallysame piston the performs four di ffperiodicallyerent strokes the(intake, four different compression, strokes expansion,(intake, exhaust),compression, in the expansion, SARM engine exhaust), each in piston the SARM performs engine continuously each piston performs the same continuously two individual the same strokes, andtwo the individual same stroke strokes, takes placeand the every same time stroke at the take sames place side. every The intaketime at process the same takes side. place The at intake the back sideprocess of the takes compression place at (CPC)the back piston, side whileof the the compression compression (CPC) is realised piston, withwhile its the front compression side. Similarly, is therealised expansion with takes its front place side. at theSimilarly, back side the of expansio the combustionn takes place (CBC) at piston,the back while side theof the burned combustion gases are exhausted(CBC) piston, to the while atmosphere the burned by gases its front are side.exhausted Thus, to all the four atmosphere strokes (intake, by its front compression, side. Thus, expansion, all four exhaust)strokes occur (intake, simultaneously compression, withexpa onension, pair exhaust) of CPC occur and CBC simultaneo pistons.usly These with mean one thatpair theof CPC CBC and piston undergoesCBC pistons. only These the high mean pressure-high that the CBC temperature piston undergoes processes, only remainingthe high pr “hot”,essure-high and thetemperature CPC piston

Energies 2020, 13, 2362 6 of 21 Energies 2020, 13, 2362 6 of 21 processes, remaining “hot”, and the CPC piston operate in the low pressure-low temperature part of the cycle, remaining “cold”. This offers an extra flexibility to the SARM engine, which is the separate operate in the low pressure-low temperature part of the cycle, remaining “cold”. This offers an extra optimization of the materials for each piston, also taking into account that they rotate in significantly flexibility to the SARM engine, which is the separate optimization of the materials for each piston, also different radii. taking into account that they rotate in significantly different radii. The concentric design is another special feature of this rotary engine. At first, the rotary operation The concentric design is another special feature of this rotary engine. At first, the rotary operation eliminates the mechanical losses imposed by the transformation of the piston’s linear motion and the eliminates the mechanical losses imposed by the transformation of the piston’s linear motion and the crankshaft’s rotational motion. In addition, one main moving part is responsible for all processes, crankshaft’s rotational motion. In addition, one main moving part is responsible for all processes, which rotates concentrically with the engine shaft. This configuration minimizes the parts of the which rotates concentrically with the engine shaft. This configuration minimizes the parts of the engine, drastically reducing its complexity and weight. On top of that, the SARM engine design has engine, drastically reducing its complexity and weight. On top of that, the SARM engine design an additional positive characteristic; the pressure force applied on the piston during combustion and has an additional positive characteristic; the pressure force applied on the piston during combustion expansion is always vertical to the piston (see Figure 3). This means that this force is always and expansion is always vertical to the piston (see Figure3). This means that this force is always transferred tangentially to the outer surface of the engine shaft and 100% of this force always transferred tangentially to the outer surface of the engine shaft and 100% of this force always produces produces torque. Contrary, in a conventional reciprocating engine, the force producing torque on the torque. Contrary, in a conventional reciprocating engine, the force producing torque on the crankshaft crankshaft is only a portion of the pressure force on the piston, depending on the instantaneous is only a portion of the pressure force on the piston, depending on the instantaneous position of the position of the crankshaft. Thus, for the same produced torque, the SARM engine needs to develop crankshaft. Thus, for the same produced torque, the SARM engine needs to develop lower combustion lower combustion , enhancing higher efficiency and allowing for lighter and thinner pressures, enhancing higher efficiency and allowing for lighter and thinner combustion chamber walls. combustion chamber walls. The higher efficiency is translated to lower fuel consumption, while the The higher efficiency is translated to lower fuel consumption, while the lower combustion pressures lower combustion pressures result to lower temperatures too, with a subsequent positive effect on result to lower temperatures too, with a subsequent positive effect on NOx formation. NOx formation.

Figure 3. The pressure force applied on the piston is always vertical. Figure 3. The pressure force applied on the 2.1.4. Advantages of the SARMpiston Engine is always vertical. Overall, the advantages of the SARM engine can be summarized as follows: Simple design and less moving parts than the reciprocating engines. Fewer engine components • reduce also the need for maintenance. Up to four times smaller and lighter in weight than the four-stroke reciprocating engines. • There are no mechanical losses and inertial forces from the conversion of reciprocating to rotational • motion. The absence of reciprocating motion eliminates oscillations [23]. No change in the direction of the piston motion, the characteristic piston “slap” occurring at the • TDC (Top Dead Center) of a conventional reciprocating engine, causing wear of the piston skirt and the bore, and noise [24]. Pressure force on the piston is always tangential and is fully exploited for power production. • In addition, the applied forces are only in the direction of the flow allowing the piston sleeve to be very short (see Figure4). Lack of lubricant inside the compression and combustion chambers. As it is well-known, • the lubricant may also contributeFigure to4. The the piston unburned sleeve HC is very emissions short of an engine [23].

Energies 2020, 13, 2362 6 of 21

processes, remaining “hot”, and the CPC piston operate in the low pressure-low temperature part of the cycle, remaining “cold”. This offers an extra flexibility to the SARM engine, which is the separate optimization of the materials for each piston, also taking into account that they rotate in significantly different radii. The concentric design is another special feature of this rotary engine. At first, the rotary operation eliminates the mechanical losses imposed by the transformation of the piston’s linear motion and the crankshaft’s rotational motion. In addition, one main moving part is responsible for all processes, which rotates concentrically with the engine shaft. This configuration minimizes the parts of the engine, drastically reducing its complexity and weight. On top of that, the SARM engine design has an additional positive characteristic; the pressure force applied on the piston during combustion and expansion is always vertical to the piston (see Figure 3). This means that this force is always transferred tangentially to the outer surface of the engine shaft and 100% of this force always produces torque. Contrary, in a conventional reciprocating engine, the force producing torque on the crankshaft is only a portion of the pressure force on the piston, depending on the instantaneous position of the crankshaft. Thus, for the same produced torque, the SARM engine needs to develop lower combustion pressures, enhancing higher efficiency and allowing for lighter and thinner combustion chamber walls. The higher efficiency is translated to lower fuel consumption, while the lower combustion pressures result to lower temperatures too, with a subsequent positive effect on NOx formation.

Energies 2020, 13, 2362 7 of 21

The symmetry of the SARM design minimizes the need for counterweights. • Up to 15% lower fuel consumption, taking advantage of the higher efficiency, as explained • previously [25]. Great potential for lower NOx emissions, due to the lower combustion temperatures. In addition, • the compression chamber can be designed in such a way that the dissipation to the coolant (or the environment) is enhanced and the charge air remains at a lower temperature (thus higher density) before entering the combustion chamber. Decoupling of the main engine processes. Although this is an inherent characteristic of any • split engine, and not only of the current one, it is important to be mentioned here that it offers great advantages compared to the conventional reciprocating piston engines. As described previously, each process occurs at a separate volume and on a specific side of each piston, enabling thermodynamic optimization,Figure 3. The pressure as well asforce selection applied ofon di thefferent materials and application of individual surface treatment.piston is always vertical.

Figure 4. The piston sleeve is very short. Figure 4. The piston sleeve is very short 2.2. Objectives and Methodology

After the brief presentation of the engine and its main technical features and advantages, this section describes the specific objectives of the current study. These are:

Investigating the most effective labyrinth seal to seal the compression chamber and • Investigate the compression chamber’s design • 2.2.1. Sealing Geometries While the SARM engine presents some particular advantages, a major concern is the sealing of the gap between its rotating arm and the engine block parts. This is a common issue in gas turbines, as well. Even though there are many advanced sealing systems developed last years, labyrinth seals remain a widely used solution for the rotating machineries [26] because they are simple, reliable, operate in high temperatures and for a wide range of pressure ratio [27]. For these reasons, this sealing method is used and optimized for the needs of the SARM engine. Labyrinth seals consist of a series of cavities connected by small clearances, avoiding the contact of the moving with the stationary parts of the sealing device (see Figure5). The high pressure of the flow is reduced due to friction through the clearance and walls, because the cavities dissipate the kinetic energy of the flow. This gradual pressure reduction continues at each cavity until the exit of the flow through the last one. Key operating parameters are the pressure ratio between the inlet and outlet of the flow from the sealing system, and the clearance (gap) between a moving tooth and the stationary wall above this tooth. Energies 2020, 13, 2362 8 of 21

For the purposes of the current study, the following labyrinth geometries are investigated (Figure 5): One-sided straight labyrinth seal (Figure 5a) where all teeth are located on the rotating part (motion transfer arm). This kind of seal is very simple to manufacture and very cheap because conventional materials may be used for its construction. The first designs of such labyrinths used large chambers between two neighbouring teeth. Moreover, relatively long teeth can be easily damaged, destroying the chambers (or pockets) between them. Tooth interlocking labyrinth seal (Figure 5b). This seal usually employs a labyrinth with more teeth and a higher pressure drop. This kind of labyrinth sealing has not been investigated thoroughly by many researchers [28,29]. For this kind of labyrinth, the size of the teeth is also investigated, which practically modifies the size of the chamber between the teeth. The following teeth dimensions are investigated: a. 3 mm at the teeth basis and 2 mm at their edge b. 2 mm at the teeth basis and 1 mm at their edge c. 1 mm at the teeth basis and 1 mm at their edge

Energies 2020 , 13Stepped, 2362 labyrinth (Figure 5c). These seals are advantageous when the clearance cannot8 of be 21 small [30]. For instance, this kind of seal is used in the compressor eye seals [31].

(a) One-sided straight labyrinth seal. (b) Tooth interlocking labyrinth seal

(c) Stepped labyrinth

Figure 5. Labyrinth geometries examined. Figure 5. Labyrinth geometries examined. For the purposes of the current study, the following labyrinth geometries are investigated (Figure5): Mesh Description 1. One-sided straight labyrinth seal (Figure5a) where all teeth are located on the rotating part (motionThe first transfer step after arm). the Thisdescription kind of of seal the is geometri very simplees is tothe manufacture construction and of the very final cheap mesh because that will conventionalbe used for the materials CFD simulations. may be used Even for its though construction. the final The mesh first is designs created of automatically such labyrinths by usedthe CONVERGElarge chambers CFD solver, between the two user neighbouring has to define teeth. some Moreover,parameters relatively for the mesh long density. teeth can The be easilyuser damaged, destroying the chambers (or pockets) between them. 2. Tooth interlocking labyrinth seal (Figure5b). This seal usually employs a labyrinth with more teeth and a higher pressure drop. This kind of labyrinth sealing has not been investigated thoroughly by many researchers [28,29]. For this kind of labyrinth, the size of the teeth is also investigated, which practically modifies the size of the chamber between the teeth. The following teeth dimensions are investigated:

a. 3 mm at the teeth basis and 2 mm at their edge b. 2 mm at the teeth basis and 1 mm at their edge c. 1 mm at the teeth basis and 1 mm at their edge 3. Stepped labyrinth (Figure5c). These seals are advantageous when the clearance cannot be small [30]. For instance, this kind of seal is used in the compressor eye seals [31].

Mesh Description The first step after the description of the geometries is the construction of the final mesh that will be used for the CFD simulations. Even though the final mesh is created automatically by the CONVERGE CFD solver, the user has to define some parameters for the mesh density. The user can Energies 2020, 13, 2362 9 of 21 Energies 2020, 13, 2362 9 of 21

can separate the model in different regions or tell the solver to refine the mesh near specific separate the model in different regions or tell the solver to refine the mesh near specific boundary boundary conditions (in this case, walls). These parameters are used for the grid-independence conditions (in this case, walls). These parameters are used for the grid-independence investigation of investigation of the solution. the solution. If the effect of rotation is excluded, then a 2D rig can provide similar results with the ones If the effect of rotation is excluded, then a 2D rig can provide similar results with the ones from from an axisymmetric 3D rig [32]. In gas turbines, like in any rotating device, the one seal side is an axisymmetric 3D rig [32]. In gas turbines, like in any rotating device, the one seal side is rotating rotating and the other is stationary. It has been found that the rotation effect becomes significant and the other is stationary. It has been found that the rotation effect becomes significant only when only when the rotating speed becomes very high. More precisely, the influencing factor is the ratio the rotating speed becomes very high. More precisely, the influencing factor is the ratio between the between the teeth’s circumferential speed and the speed of the flow. When this is very high, then teeth’s circumferential speed and the speed of the flow. When this is very high, then the rotation effect the rotation effect becomes significant [33]. Therefore, static rigs are mostly preferred for this becomes significant [33]. Therefore, static rigs are mostly preferred for this reason [30]. reason [30]. Not only the turbulence model introduces an error, but also two additional effects influence the Not only the turbulence model introduces an error, but also two additional effects influence results’ accuracy. The first is the discretization parameter, which affects the prediction accuracy when the results’ accuracy. The first is the discretization parameter, which affects the prediction accuracy coarse grids are used. In order to minimize this error, after running some simulations, it was concluded when coarse grids are used. In order to minimize this error, after running some simulations, it was that using 1,000,000 cells is the best approach. This is also proven by the results in Figure6, which concluded that using 1,000,000 cells is the best approach. This is also proven by the results in Figure presents the influence of the grid resolution on the discharge coefficient (DC) and mass flow rate for the 6, which presents the influence of the grid resolution on the discharge coefficient (DC) and mass case of one-sided labyrinth. The same investigation concerning grid independency was conducted for flow rate for the case of one-sided labyrinth. The same investigation concerning grid independency all the geometries considered in this study, concluding that the accuracy of the results was acceptable was conducted for all the geometries considered in this study, concluding that the accuracy of the when using meshes of more than 1,000,000 cells. results was acceptable when using meshes of more than 1,000,000 cells.

0.678 0.7 0.64 0.68 0.65 0.66 0.6 0.59 0.55 0.477 0.5 0.48 0.45 0.45 0.46 0.4 0.4 0.35 0.3 0 1000 2000 3000 4000 Number of quads (in thousands)

flow rate (kg/s) CD

Figure 6. How the discharge coefficient (DC) and the mass flow rate changes with the grid resolution. Figure 6. How the discharge coefficient (DC) and the mass flow rate changes with the grid Theresolution. second influence comes from the boundary layer of the wall. The initial assumption was that the high pressures would develop velocities that are too high, thus the influence of the boundary layer wouldThe be negligible.second influence However, comes a sensitivity from the analysisboundary conducted layer of showedthe wall. that The the initial boundary assumption layer has was a significantthat the high eff pressuresect. Particularly would develop in the case velocities of a very that fine are grid, too high, the closest-to-the-wall thus the influence of grid the point boundary falls belowlayer would values be of negligible. y+ = 30. In However, this region, a sensitivity the logarithmic analysis law conducted is not valid showed {7}. Anthat ANSA the boundary (BETA CAElayer Systems)has a significant pre-processor effect. wasParticularly used for in the the creation case of of a thevery geometries. fine grid, the closest-to-the-wall grid point falls below values of y+ = 30. In this region, the logarithmic law is not valid {7}. An ANSA (BETA CAE CFDSystems) Model pre-processor Setup was used for the creation of the geometries. The numerical simulations of the labyrinth seal flow for the different geometries were executed CFD Model Setup with the commercial software CONVERGECFD, solving the time-averaged Navier–Stokes equations. The advantageThe numerical of this simulations software is theof the fast labyrinth setup and seal solving flow offor many the different complicated geometries design configurationswere executed becausewith the the commercial mesh is created software automatically. CONVERGECFD, solving the time-averaged Navier–Stokes equations. The Inadvantage the current of investigation, this software the wallsis the were fast considered setup and smooth solving and withof many zero slip.complicated This assumption design wasconfigurations made in order because to simplify the mesh somehow is created the automatically. problem and facilitate the faster convergence of the code. It is acknowledged,In the current however,investigation, that the the wall walls roughness were considered would increase smooth the circumferentialand with zero velocityslip. This of theassumption fluid before was entering made in the order seal. to simplify somehow the problem and facilitate the faster convergence of the code. It is acknowledged, however, that the wall roughness would increase the circumferential velocity of the fluid before entering the seal.

Energies 2020, 13, 2362 10 of 21

Another simplification of the problem came from the assumption that the working fluid is air, which is assumed as an , in a steady and adiabatic flow. This approach is a common approach for labyrinth seal investigations [34–36]. The parameter used to confirm convergence was the Root Mean Square residuals (RMS) of the momentum, energy, mass and turbulence equations and the limit value applied as convergence 6 criterion was set to 10− . Turbulence models and the relevant assumptions are summarized in Table1.

Table 1. CFD model setup.

Setup Selected options Solver Navier-Stokes Pressure-Based Convergent Criterion PISO 3D Periodic Boundary/Sector (pie shape) Flow Assumption Steady, ideal gas compressible, turbulent Models Time based, SST k-ω Materials Air (ideal gas) Operating Pressure (bar) 0 Pressure-inlet (bar) 50 Inlet Temperature (K) 700 Pressure-outlet (bar) 2 Outlet Temperature (K) 450 Wall properties Adiabatic, absolute roughness=0, no-slip Solution Methods Scheme Transient Spatial Discretization Second Order Upwind Solution Initialization Hybrid 6 Convergence criterion PISO tolerance 10−

As mentioned above, the flow was considered adiabatic. This assumption is based on the study of Mirzamoghadam and Xiao [37] who conducted a numerical study on an industrial gas turbine with labyrinth seal. They compared the temperature and pressure in the labyrinth, for an adiabatic wall and a wall with heat transfer. They concluded that the profiles do not change with the heat transfer. They verified their conclusion also with steady-state temperature and static pressure measurements in the labyrinth. Table1 summarizes all parameters and values used for the CFD model setup. Thorat and Chids showed that the whirl frequency depends significantly on the DC of the labyrinth seals when the Mach number approaches unity [38], or when there is an inlet pre-swirl or when the pressure is very high [38,39]. Additionally, pre-swirl and the rotor whirl speeds have a small effect on the leakage for one-sided and stepped labyrinths [29]. The SST k-ω equations were used for the turbulence model. The effects were considered as Mach number approaches 1.0. The SST k-ω model was chosen over k-ε, as it is more used for complex boundary layer flows with adverse pressure gradient and separation. This is very common for external aerodynamics and turbomachinery. 6 As mentioned above, the desired convergence target was set to 10− (RMS residuals for the momentum, mass, energy, and turbulence equations). The net flux imbalance was less than 1% of the smallest flux in the whole boundary domain for all converged cases. Since DC and pressure ratio will present all calculation results, the setup of the inlet and outlet boundary conditions does not matter significantly [30]. The good setup of the solver is verified by the fast convergence of all solution-cases.

2.2.2. Compression Chamber’s Design

Simulation Model Setup The first step is the 3D CAD design of the volume included inside the compression chamber and the labyrinth, which is prepared in ANSA (BETA CAE Systems). The major parameter to begin Energies 2020, 13, 2362 11 of 21 with is the size of the toroidal compression chamber, quantified by its outer diameter (Rcpc, Figure3). The principle idea is to keep the compression chamber as close as possible to the engine shaft (i.e., small Rcpc). However, the sealing system limits the minimum possible value of this diameter because there is a minimum length of each labyrinth seal type that makes the labyrinth effective (Lab.length, Figure3). The labyrinth must be long enough to prevent any leakage of the working medium from the chamber to the environment. The above study tests the labyrinth in the most extreme thermodynamic conditions developed inside the combustion chamber. The compression chamber’s design study tests only the two most effective labyrinths for sealing the compression chamber from the environment; the stepped labyrinth seal and the tooth-interlocking or two-sided labyrinth (Figure5). The stepped labyrinth needs a shorter length to drop down the pressure, while the two-sided labyrinth needs a longer path. Therefore, the stepped labyrinth is characterised in the following analysis as the Long-Sealing system (LS) and the two-sided labyrinth as Short-Sealing (SS). Even though the most logical thought is to use the stepped labyrinth, since it is a shorter (more efficeint) labyrinth, this labyrinth needs a thicker body for the motion arm and the housing due to the steps (Figure5c). The longer the stepped labyrinth, the higher the first tooth from the last one; this is translated to a thicker motion disc (Figure3). On the other hand, the two-sided labyrinth is longer but its length does not influence the thickness of the motion disc. Like Figure5a shows, all teeth are on the same height-level. After defining the 3D fluid-volume geometry inside ANSA, follows the definition of the boundary conditions. The fluid-volume is a closed volume, defined by the walls of the housing, sliding port, piston and labyrinth, as well as the inlet boundary on the intake port and the outlet boundary at the end of the labyrinth. Here, the only assumption is that the sliding port seals watertight when it closes. When the sliding port opens, a very small percentage of the trapped air leaves the chamber (10%) through the intake port, but eventually comes back when the intake process starts again. The complexity of the fluid’s design (chamber with piston, sliding port and labyrinth) does not allow for the creation of an understandable 3D model view. On the other hand, ANSA topology tools offer an easy and fast preparation of the final watertight 3D model. Before exporting the STL file to the solver, the minimum number of nodes is essential to keep the final file as small as possible, for making the automatic procedure of mesh generation inside the solver faster. The CONVERGE CFD solver creates both the surface and volume mesh automatically and the only requirement is to input the STL file. It is crucial to define all 3D geometrical details with the minimum number of nodes. A very-fine STL mesh avoids intersections of the parts, especially when there are many curvatures in the geometry. On the other hand, a very fine mesh in regions where there are only flat surfaces can make the STL file heavy which makes the simulation process slower; the higher the number of nodes, the slower the generation of the mesh during each time-step. Since CONVERGE CFD can read only STL files to recognise the model’s geometry, the model has to be initially meshed inside ANSA with the STL mesh algorithm. The meshing tools of ANSA allow for the automatic creation of an STL mesh with the least possible nodes in flat surfaces and a controllable number of nodes in the curved edges. The final quality-control is to smooth all curved surfaces with the ortho-triangle mesh. This is an automatic procedure with no additional time required by the user. CONVERGE CFD is mainly designed for simulations in internal flows and thermal engines of any nature, such as reciprocating or rotary ones, turbomachines, pumps and compressors, boilers and burners. It is ideal for highly transient simulations with a variable mesh. Its main advantage compared to other CFD solvers is the automatic creation of the grid. This solver creates a perfect rectangular structured grid based on simple user-defined parameters [40] (see Figure7). Moreover, the user can use a rotation angle instead of time, which is ideal for rotational devices like the SARM engine. EnergiesEnergies 20202020,, 1313,, 23622362 12 of 21 Energies 2020, 13, 2362 12 of 21

Figure 7. the mesh created by the “CONVERGE CFD“ solver for the stepped labyrinth seal. FigureFigure 7. the 7. The mesh mesh created created by the by the“CONVERGE “CONVERGE CFD“ CFD“ solver solver for forthe thestepped stepped labyrinth labyrinth seal. seal. The turbulence model used for this compressible flow is k-e RNG. Although the RSM model is The turbulence model used for this compressible flow is k-e RNG. Although the RSM model idealThe for turbulencesimulating modelflows usedwith forintense this compressiblevortices, pressure flow iswaves, k-e RNG. high Although turbulence the and RSM directional model is is ideal for simulating flows with intense vortices, pressure waves, high turbulence and directional idealchanges for (suchsimulating as curved flows pipes, with cyclones), intense vortices, a compromise pressure was waves, necessary high between turbulence the andsimulation directional time changes (such as curved pipes, cyclones), a compromise was necessary between the simulation time changesand accuracy (such of as the curved results pipes, due cyclones),to the high a computationalcompromise was cost. necessary The k-e RNGbetween model the issimulation widespread time in and accuracy of the results due to the high computational cost. The k-e RNG model is widespread in andfour-stroke accuracy Internal of the results Combustion due to Enginesthe high (ICE),computational and after cost.a benchmark The k-e RNG between model the is two widespread turbulence in four-stroke Internal Combustion Engines (ICE), and after a benchmark between the two turbulence four-strokemodels, the Internal results showedCombustion that thereEngines is a (ICE),small anddifference after a in benchmark the solution, between but with the an two exceptionally turbulence models, the results showed that there is a small difference in the solution, but with an exceptionally models,longer simulation-time the results showed for the that RSM there [41]. is aFinally, small di thefference META in post-processor the solution, butwas with used an for exceptionally the analysis. longer simulation-time for the RSM [41]. Finally, the META post-processor was used for the analysis. longer simulation-time for the RSM [41]. Finally, the META post-processor was used for the analysis. Model Optimization Model Optimization Before runningrunning anyany case,case, thethe time,time, gridgrid andand iterationiteration cyclecycle independencyindependency isis assessed.assessed. According toto CONVERGECONVERGEBefore running CFDCFD any recommendations, recommendations, case, the time, grid for for ICEs and ICEs ititeration isit essentialis essential cycle first independency first to runto run at least at isleast 5assessed. to 5 30 to iterations 30 According iterations [40] toin[40] orderCONVERGE in order to obtain to obtainCFD a solution recommendations, a solution that isthat not is a notff forected affected ICEs by it the isby operatingessential the operating first cycles. to cycles. run at least 5 to 30 iterations [40] inBeginning order to obtain with the a solution time step, that thethe is solversolvernot affected ensuresensure bys a athe time-independenttime-independent operating cycles. study by usingusing aa variablevariable 8 3 time-step.time-step.Beginning By choosingchoos withing the aa minimumtimeminimum step, oftheof 1010 solver−-8 andand ensurea a maximum maximums a time-independent of 10 -3− s stime-step, time-step, study for for 3,000 3000by using rpm, athe variable solver time-step.uses the optimum By choos valuevalueing a minimum accordingaccording oftoto 10thethe-8 movingandmoving a maximum meshmesh velocityvelocity of 10-3 andsand time-step, pressurepressure for changechange 3,000 rate.rpm,rate. the solver uses Forthe optimum the grid size value eeffect,ff ect,according itit waswas foundfoundto the thatmovingthat the resultmeshresult velocity usingusing a gridand pressureof 400,000 change cells deviates rate. only by lessless thanthanFor the 2%2% grid fromfrom size thethe effect, finestfinest it gridgrid was (of(of found 1,100,0001,100,000 that the cells).cells). resu Atlt using the same a grid time, of 400,000 this coarse cells grid deviates requires only onlyonly by less40% than of the 2% computationalcomputational from the finest timetime grid ofof (of thethe 1,100,000 finestfinest grid.grid. cells). At the same time, this coarse grid requires only 40% ofAfter the choosingchoosingcomputational thethe rightright time time-step time-step of the finest and and grid-size,grid. grid-size, the the model model is is executed executed iteratively iteratively for for anumber a number of cycles.of cycles.After As As choosing presented presented the in in Figureright Figure time-step8, 8, after after four andfour cycles grid-size, cycles the the deviation the deviation model in inis the executedthe maximum maximum iteratively calculated calculated for a pressure pressurenumber ofand cycles. temperaturetemperature As presented drops in belowbelow Figure 1%1% 8, after betweenbetween four twocycltwo es consecutiveconsecutive the deviation cycles.cycles. in the Thus, maximum there wascalculated no need pressure to runrun andmore temperature thanthan fivefive cycles.cycles. drops below 1% between two consecutive cycles. Thus, there was no need to run more than five cycles.

pressure temperature pressure temperature 39% 39% 28% 19% 28% 19% 5.50% -1% 1% 2% 0.80% -1% 5.50% -0.50% -1% 1% 2%-0.90% 0.80% -1% 1234-0.90% -0.50% 1234CYCLE NUMBER

PERCENTAGE CHANGE CYCLE NUMBER PERCENTAGE CHANGE Figure 8. EffectEffect of the engine rotations’ number on the maximum temperature and pressure showing Figurethethe cyclecycle 8. independency.Effectindependency. of the engine rotations’ number on the maximum temperature and pressure showing the cycle independency.

Energies 2020, 13, 2362 13 of 21 Energies 2020, 13, 2362 13 of 21 3. Results and Discussion 3. Results and Discussion 3.1. Sealing Optimization 3.1. Sealing Optimization Pressure contour plots were produced for all the different labyrinth geometries. According to previousPressure research contour [30], when plots werethe clearance produced is forsuffic alliently the di ffsmall,erent labyrinththen the pressure geometries. distribution According is to uniformprevious along research the cavities. [30], when On the the clearance other hand, is su inffi cientlythe case small, of a larger then the clearance, pressure the distribution pressure isdrop uniform is relativelyalong the greater cavities. in Onthe thefirst other cavity. hand, In inthe the geom caseetries of a larger examined clearance, in this the pressurestudy, the drop clearance is relatively is sufficientlygreater in small the first (200 cavity. μm), thus In the the geometries uniform pressure examined drop in thisis expected, study, the validating clearance the is sufinalfficiently results small of the(200 CFDµm), model. thus the uniform pressure drop is expected, validating the final results of the CFD model. TheThe following following figures figures present present the the gas gas velocity velocity co contoursntours inside inside the the different different seal seal geometries: geometries: • FigureFigure 9: one-sided9: one-sided straight straight labyrinth labyrinth • • FigureFigure 10: 10tooth-interlocking: tooth-interlocking labyrinth labyrinth with with1 mm 1 (top), mm (top),2 mm 2(middle) mm (middle) and 3 mm and (bottom) 3 mm (bottom) basis • thicknessbasis thickness • FigureFigure 11: 11stepped: stepped labyrinth labyrinth • In Inall all cases, cases, the the CFD CFD model model converged converged according according to tothe the criteria criteria set set in inthe the previous previous section, section, producingproducing the the streamlines streamlines insideinside the the chambers chambers of theof dithefferent different labyrinth labyrinth types. Thetypes. various The geometriesvarious geometriesgenerate digeneratefferent flowdifferent fields, flow while fields, in all while the cases in all the the working cases the fluid working exits finallyfluid exits from finally the tooth from at a thehigh tooth velocity. at a high For eachvelocity. individual For each geometry, individual the flow geometry, field is similar the flow among field the is intermediate similar among chambers. the intermediateIn the chambers. cases of the one-sided straight labyrinth (Figure9), the tooth-interlocking labyrinth (FigureIn the 10 cases) and of the the stepped one-sided labyrinth straight (Figure labyrinth 11 ),(Figure the high 9), velocitythe tooth-interlocking streamlines are labyrinth located (Figure at the top 10)side and of the the stepped last chamber. labyrinth (Figure 11), the high velocity streamlines are located at the top side of the lastIn chamber. the case of the stepped seal, the large step distance combined with the sufficient step height increasesIn the case the dissipated of the stepped dynamic seal, energy the large inside step the di cavity,stance ascombined compared with to anythe sealsufficient with straightstep height shape. increasesThis complicated the dissipated flow dynamic structure energy incurs ainside larger th pressuree cavity, dropas compared for a given to massany seal flow. with On straight the other shape.hand, This a disadvantage complicated of flow the steppedstructure seal incurs is its a largerlarger overallpressure seal drop height. for a Hence, given themass overall flow. sizeOn ofthe the otherstepped hand, seal a disadvantage is larger than of that the ofstepped any straight seal is seal.its larger overall seal height. Hence, the overall size of the stepped seal is larger than that of any straight seal.

Figure 9. Velocity contours inside the one-sided straight labyrinth. Figure 9. Velocity contours inside the one-sided straight labyrinth

Energies 2020, 13, 2362 14 of 21 Energies 2020, 13, 2362 14 of 21

(a)

(b)

Figure 10. Cont.

Energies 2020, 13, 2362 15 of 21 Energies 2020, 13, 2362 15 of 21 Energies 2020, 13, 2362 15 of 21

(c) (c) Figure 10. Velocity contours inside the tooth interlocking labyrinth with 1 mm (a), 2 mm (b) and 3 FigureFigure 10. 10.Velocity Velocity contours contours inside inside the the tooth tooth interlocking interlocking labyrinth labyrinth with with 1 1 mm mm ( a(),a), 2 2 mm mm (b ()b and) and 3 3 mm mm (c) basis thickness. (c)mm basis (c) thickness. basis thickness.

Figure 11. Velocity contours inside the stepped labyrinth FigureFigure 11. 11. VelocityVelocity contours contours inside inside the the stepped stepped labyrinth labyrinth. In order to assess comparatively the different labyrinth geometries, and since such a direct In order to assess comparatively the different labyrinth geometries, and since such a direct comparisonIn order to is assess not straight-forward comparatively with the the different velocity labyrinth contour plots, geometries, the pressure and drop since across such the a directseals comparison is not straight-forward with the velocity contour plots, the pressure drop across the seals comparisonis examined, is not as straight-forward shown in Figure with12. A the significant velocity pressure contour dropplots, is the observed pressure exactly drop after across each the tooth seals is isexamined,tip. examined, Inside as each asshown shown cavity, in inFigurethe Figure pressure 12. 12 A. remains Asignificant significant almost pressure pressure constant drop dropand is a isobserved small observed pressure exactly exactly rise after is after observed each each tooth toothat tip.tip. Insidethe Inside leading each each edgecavity, cavity, of eachthe the pressure tooth. pressure The remains remainslatter is almost the almost resu constantlt constant of the andpartial and a small a blocking small pressure pressure of the rise flow rise is at isobserved the observed tooth, at at thethe leadingreversing leading edge the edge flowof ofeach eachback tooth. to tooth. the The chamber. The latter latter is the is the resu resultlt of ofthe the partial partial blocking blocking of ofthe the flow flow at atthe the tooth, tooth, reversingreversing the the flow flow back back to tothe the chamber. chamber.

Energies 2020, 13, 2362 16 of 21

For the quantification of the sealing efficiency, the (dimensionless) DC is determined. This coefficient expresses the fraction of the leakage mass flow to the ideal mass flow rate for an isentropic expansion through a single nozzle [42]. This means that the target is to obtain a low value for this coefficient. As shown in Figure 12, the average DC of the stepped seal is lower than that of most of the straight seals, indicating that the leakage performance of this geometry seal is better. However, this type of labyrinth is characterized by a more expensive construction, as well as by larger overall height. The above observations indicate clearly that for the specific operating conditions (pressure ratio 25 and temperature 700 K), the stepped labyrinth has the best efficiency, with a DC equal to 0.38 and the lowest mass flow rate. On top of this, according to Waschka et al. [33], the rotation results in a Energies 2020, 13, 2362 16 of 21 further reduction of the DC, when the circumferential velocity of the rotor exceeds the axial velocity of the flow. This means that at high engine speeds, the sealing efficiency is expected to be even higher.

1.000 0.900 0.784 0.800 0.700 0.627 0.636 0.611 0.600 0.500 0.393 0.400 0.351 0.281 0.285 0.273 0.300 0.176 0.200 0.100 0.000 one-sided two-sided two-sided two-sided stepped (3mm) (2mm) (1mm)

mass flow (kg/s) CD

FigureFigure 12.12. InfluenceInfluence ofof geometrygeometry configurationconfiguration onon thethe computedcomputed dischargedischarge coecoefficientfficient andand thethe massmass flowflow rate.rate. For the quantification of the sealing efficiency, the (dimensionless) DC is determined. This 3.2. Compression Chamber Configuration coefficient expresses the fraction of the leakage mass flow to the ideal mass flow rate for an isentropic expansionThe first through output a single of the nozzle model [42 ].is Thisthe examination means that theof targetthe compression is to obtain achamber low value (CPC) for thisconfiguration. coefficient. According to previous work [20], the target is to keep the CPC as close as possible to the engineAs shown shaft. in FigureHowever, 12, thethe averagetwo different DC of sealin the steppedg systems seal isrequire lower a than different that of Rcpc most (Figure of the straight 3). The seals,stepped indicating labyrinth that (Short-Seal: the leakage SS) performance uses the mi ofnimum this geometry possible seal Rcpc is better.and the However, two-sided this (LS) type the of labyrinthmaximum is possible. characterized However, by a morethe Δ expensiveRcpc (Rcpcmax-Rcpcmin) construction, as wellis limited as by largerto 10mm overall due height.to the under constraintsThe above of the observations engine’s indicatedesign for clearly keeping that forthe theoverall specific size operating within specific conditions limits. (pressure The outer ratio 25diameter and temperature of the engine 700 cannot K), the exceed stepped the labyrinth 200mm in has order the bestto use effi CNCciency, machining with a DC with equal low to cost. 0.38 Here, and thethe lowestcomparison mass of flow the rate.two chamber’s On top of configurations this, according is to based Waschka on the et final al. [33 peak], the pressure rotation at results the end in of a furtherthe compression reduction process of the DC, (Table when 2). the circumferential velocity of the rotor exceeds the axial velocity of the flow.The Thissecond means output that of at the high model engine concerns speeds, thethe sealingoptimization efficiency of the is expected shape of to the be intake even higher. port in order to maximise the volumetric efficiency. Two intake port designs are considered: a rectangular 3.2.(i.e., Compression the Initial Intake, Chamber II) Configurationand another one that offers the optimal shape for the most efficient charge of theThe compression first output chamber of the model with isfresh the examinationair (i.e., the Optimized of the compression Intake, OI). chamber (CPC) configuration. AccordingTable to2 presents previous the work results [20 ],of the the targetCFD calculatio is to keepns the and CPC the asones close assuming as possible isentropic to the process. engine shaft.Initially, However, the sealing the system two di ffiserent assessed sealing (LS systemsvs. SS) and require the best a di isff simulatederent Rcpc with (Figure the 3optimized). The stepped intake labyrinthport (OI). (Short-Seal: SS) uses the minimum possible Rcpc and the two-sided (LS) the maximum possible. However, the ∆Rcpc (Rcpcmax-Rcpcmin) is limited to 10mm due to the under constraints of the engine’s design for keeping the overall size within specific limits. The outer diameter of the engine cannot exceed the 200mm in order to use CNC machining with low cost. Here, the comparison of the two chamber’s configurations is based on the final peak pressure at the end of the compression process (Table2).

Table 2. Optimization of intake port and sealing system.

CFD Configuration Peak p (bar) Peak T (K) Peak p (bar) Peak T (K) LS.II 23.3 587 28.9 775 SS.II 25.6 617 28.9 775 SS.OI 26.9 614 28.9 775 Energies 2020, 13, 2362 17 of 21

Energies 2020, 13, 2362 Table 2. Optimization of intake port and sealing system. 17 of 21

Energies 2020, 13, 2362 CFD Isentropic process17 of 21 Configuration The second outputPeak of p the (bar model) concernsPeak the T optimization (K) ofPeak the shapep (bar) of the intakePeak port T (K in) orderLS.II to maximise the volumetricTable23.3 2. Optimization efficiency. of Two intake 587 intake port portand sealing designs system. are 28.9 considered: a rectangular775 (i.e.,SS.II the Initial Intake, II) and25.6 another oneCFD that offers 617 the optimal shape for 28.9Isentropic the most eprocessfficient charge775 of Configuration the compressionSS.OI chamberPeak26.9 with p ( bar fresh) air (i.e., thePeak Optimized 614 T (K ) Intake,Peak OI). p 28.9 (bar ) Peak T775 (K) TableLS.II2 presents the results23.3 of the CFD calculations 587 and the ones 28.9 assuming isentropic775 process. As shown in Figure 13, the combination of the optimized intake-port and the short sealing Initially,SS.II the sealing system25.6 is assessed (LS vs. SS) 617 and the best is simulated 28.9 with the optimized775 intake system can reach up to 15% higher peak pressure, as compared to the initial configuration. port (OI).SS.OI 26.9 614 28.9 775 As shown in Figure 13, the combination of the optimized intake-port and the short sealing system As shown in Figure 13, the combination of the optimized intake-port and the short sealing can reach28 up to 15% higher peak pressure, as compared to the initial configuration. system can reach up to 15% higher peak pressure, as compared to the initial configuration. 27 26 5% 28 25 27 15% 10% 24 26 5% 2325 15% 10% peak pressure (bar)peak pressure 2224 2123 L.S.II SS.II SS.OI peak pressure (bar)peak pressure 22 21 FigureL.S.II 13. Optimization of intake SS.IIport shape and sealing length. SS.OI

The following figures present the development of pressure and temperature in the intake and FigureFigure 13. 13.Optimization Optimization of intake port shape shape and and sealing sealing length. length. compression chambers. Beginning with the intake chamber, Figure 14 highlights the advantageous configurationTheThe following following of the figures figures optimized present present intake thethe development developmentport combin edof of pressure pressurewith the and andshort temperature temperature sealing system in inthe the intake(SS.OI). intake and andTwo compressionpositivecompression aspects chambers. chambers. are revealed. Beginning Beginning On the withwith one the hand, intake the ch chamber, amber,SS.OI configuration Figure Figure 14 14 highlights highlights results the in the advantageoushigher advantageous pressure, configurationalmostconfiguration throughout of theof the optimizedthe optimized complete intake intake port portprocess. combined combin Th withised enhances with the shortthe theshort sealing charging sealing system ofsystem the (SS.OI). CPC (SS.OI).Two with positive Twoair and aspectsleadspositive to are a aspectshigher revealed. arepeak revealed. On pressure the one On at hand,the the one end the hand, of SS.OI compression. the configuration SS.OI configuration On the results other results in hand, higher in thehigher pressure, pressure pressure, almost in the throughoutintakealmost chamber throughout the complete with the the complete intake SS.OI process. intakeconfiguration process. This enhances This isal enhancesmost the constant charging the charging throughout of the of CPC the CPCthe with intakewith air andair process, and leads tominimizingleads a higher to a peak higherany pressurenegative peak pressure flow at the effectsend at the of due end compression. to of pressu compression.re Ongradient. the On other the The other hand, lower hand, the pressure pressure the pressure at the in thebeginning in intakethe intake chamber with the SS.OI configuration is almost constant throughout the intake process, chamberof the intake with process the SS.OI with configuration the other two is configuratio almost constantns (both throughout with the theinitial intake intake process, port shape) minimizing shows minimizing any negative flow effects due to pressure gradient. The lower pressure at the beginning anyclearly negative that the flow initial effects intake due port to pressure design causes gradient. a pressure The lower drop. pressure Concerning at the temperature beginning of in the the intake intake of the intake process with the other two configurations (both with the initial intake port shape) shows processchamber, with there the is other only twoa slight configurations difference between (both with the the different initial intakeconfigurations. port shape) shows clearly that clearly that the initial intake port design causes a pressure drop. Concerning temperature in the intake the initialPressure intake and port temperature design causes development a pressure in drop. the co Concerningmpression chamber temperature are presented in the intake in Figure chamber, 15. Thechamber, differences there isbetween only a slight the threedifferen configurationsce between the examineddifferent configurations. here are minimal with an obvious there isPressure only a slight and temperature difference between development the di inff erentthe co configurations.mpression chamber are presented in Figure 15. difference only on the peak pressure and temperature, as already shown in Table 2. The differences between the three configurations examined here are minimal with an obvious difference only on the peak pressure and temperature, as already shown in Table 2.

(a) (b) (a) (b) FigureFigure 14.14. PressurePressure ((aa)) andand temperaturetemperature ((bb)) developmentdevelopment inin thethe intakeintake chamber. chamber. Figure 14. Pressure (a) and temperature (b) development in the intake chamber.

Energies 2020, 13, 2362 18 of 21

Pressure and temperature development in the compression chamber are presented in Figure 15. The differences between the three configurations examined here are minimal with an obvious difference onlyEnergies on 2020 the, peak13, 2362 pressure and temperature, as already shown in Table2. 18 of 21

(a) (b)

FigureFigure 15.15. PressurePressure ((aa)) andand temperaturetemperature (b) development in in the the compression compression chamber. chamber.

4.4. Conclusions Conclusions TheThe current current workwork studiesstudies twotwo particularparticular featuresfeatures of aa novelnovel rotary engine. The The SARM SARM engine engine presentspresents a a number number of of advantages advantages asas comparedcompared toto conventionalconventional reciprocating piston engines. engines. The The most most importantimportant advantages advantages are are its its simple simple design design and and the the fewer fewer moving moving parts parts that that are are directly directly translated translated to smallerto smaller and lighterand lighter construction. construction. The mainThe dimainfference difference from otherfrom rotary other enginesrotary engines is that the is processesthat the areprocesses decoupled are decoupled and are realized and are in realized different in chambers,different chambers, allowing allowing for individual for individual optimization. optimization. Possible advantagesPossible advantages are lower are fuel lower consumption fuel consumption and great and potential great forpotential significantly for significantly reduced NOx reduced emissions. NOx emissions.In the present study, the compression chamber design and the sealing system were investigated. For thatIn the aim, present simulation study, models the compression were setup chamber in CONVERGECFD design and the software. sealing system The required were investigated. meshes for solvingFor that the aim, momentum, simulation mass, models energy were and setup turbulence in CONVERGECFD equations were software. constructed The required automatically meshes by for the softwaresolving the and momentum, the relevant mass, assumptions energy and were turbulence made. equations were constructed automatically by the softwareTwo parameters and the ofrelevant the compression assumptions chamber were made. were examined, namely the intake port design and the lengthTwo ofparameters the sealing of system.the compression The initial chamber and the were optimized examined, design namely were the examined intake port for design the former, and whilethe length a short of andthe sealing a long sealingsystem. system The initial were and considered. the optimized design were examined for the former, whileAt a first, short the and compression a long sealing chamber system (CPC) were considered. configuration was investigated. Previous work shows that keepingAt first, the the CPC compression as close as chamber possible to(CPC) the engineconfigur shaftation is beneficialwas investigated. [20,25]. ThePrevious limiting work parameter shows ofthat the keeping minimum the rotation CPC as radius close of as the possible CPC piston, to the however, engine shaft is the is sealing beneficial system. [20,25]. Then, The with limiting the use ofparameter the CFD of model, the minimum it was found rotation that radius the optimum of the CPC design piston, of thehowever, CPC is is with the sealing the optimized system. Then, intake portwith and the the use short of the seal. CFD With model, this configuration, it was found athat higher the peakoptimum pressure design is achieved of the CPC at the is endwith of the the compressionoptimized intake stroke. port and the short seal. With this configuration, a higher peak pressure is achieved at theThe end second of the featurecompression of the stroke. engine examined in this study was the sealing system, where three differentThe labyrinth second feature geometries of the were engine considered, examined namely in this the study one-sided was the straight, sealing thesystem, tooth-interlocking where three (threedifferent variances), labyrinth and geometries the stepped were labyrinth. considered, The name steppedly the design one-sided exhibited straight, the the best tooth-interlocking sealing efficiency and(three it is variances), considered and preferable. the stepped labyrinth. The stepped design exhibited the best sealing efficiency andCurrently, it is considered and after preferable. concluding with the modeling analysis of the engine, the first prototype of this novelCurrently, engine and is underafter concluding construction. with The the initialmodeling parts analysis that are of beingthe engine, manufactured the first prototype include theof compressionthis novel engine chamber is andunder the construction. CPC pistons. The Those initial will constituteparts that theare core being of themanufactured experimental include installation the thatcompression is under development. chamber and In the parallel, CPC apistons. test campaign Those iswi designedll constitute for the the first core experimental of the experimental assessment ofinstallation the engine, that the resultsis under of whichdevelopment. will be also In usedparall forel, thea test validation campaign of the is simulationdesigned models.for the first experimental assessment of the engine, the results of which will be also used for the validation of the simulation models.

Energies 2020, 13, 2362 19 of 21

Supplementary Materials: The following are available online at http://www.mdpi.com/1996-1073/13/9/2362/s1; Video S1: Operating Principle of the SARM engine. Author Contributions: Conceptualization, S.S. and Y.Y.; methodology, S.S.; software, S.S.; validation, S.S.; formal analysis, S.S.; investigation, S.S.; resources, S.S.; data curation, S.S.; writing—original draft preparation, S.S.; writing—review and editing, E.N., D.M. and Z.S.; visualization, S.S.; supervision, Z.S.; project administration, Z.S.; funding acquisition, Z.S. All authors have read and agreed to the published version of the manuscript. Funding: This research is co-financed by Greece and the European Union (European Social Fund- ESF) through the Operational Programme “Human Resources Development, Education and Lifelong Learning 2014–2020” in the context of the project “Optimisation of the sealing method for a new concept concentric rotary engine” (MIS 5005530). Acknowledgments: We would like to thank BETA CAE Systems & Convergent Science for their free of charge offer of their software (ANSA, META & CONVERGECFD) and their technical support for the accomplishment of this research. ANSA was used for the geometry creation and pre-processing of the CFD model, CONVERGECFD for the CFD simulations, and META for the post-processing/visualization of the results. Their support was essential and crucial to complete this research. Conflicts of Interest: The authors declare no conflict of interest. The funders had no role in the design of the study; in the collection, analyses, or interpretation of data; in the writing of the manuscript, or in the decision to publish the results.

References

1. European Commission. Regulation (EU) 2019/631 of the European Parliament and of the Council of 17 April

2019, Setting CO2 Emission Performance Standards for New Passenger Cars and for New Light Commercial Vehicles, and Repealing REGULATIONS (EC) No 443/2009 and (EU) No 510/2011; European Commission: Brussels, Belgium, 2019. 2. Federal Register, Department of Transportation. 2017 and Later Model Year Light-Duty Vehicle Greenhouse Gas Emissions and Corporate Average Fuel Economy Standards; Final Rule; U.S. Environmental Protection Agency: Washington, DC, USA, 2012; Volume 77, pp. 62623–63200. 3. European Commission. Regulation (EU) 2019/1242 of the European Parliament and of the Council of 20 June 2019, Setting CO2 Emission Performance Standards for New Heavy-Duty Vehicles and Amending Regulations (EC) No 595/2009 and (EU) 2018/956 of the European Parliament and of the Council and Council Directive 96/53/EC; European Commission: Brussels, Belgium, 2019. 4. Statharas, S.; Moysoglou, Y.; Siskos, P.; Zazias, G.; Capros, P. Factors Influencing Electric Vehicle Penetration in the EU by 2030: A Model-Based Policy Assessment. Energies 2019, 12, 2739. 5. Yugo, M.; Soler, A. A Look into the Role of e-Fuels in the Transport System in Europe (2030–2050) (Literature Review). Available online: https://www.concawe.eu/wp-content/uploads/E-fuels-article.pdf (accessed on 1 April 2020). 6. Ash, N.; Sikora, I.; Richelle, B. Electrofuels for Shipping: How Synthetic Fuels from Renewable Electricity Could Unlock Sustainable Investment in Countries Like Chile; Environmental Defense Fund: London, UK, 2019. 7. ERTRAC (European Road Transport Research Advisory Council) Working Group: Energy and Environment. Future Light and Heavy Duty ICE Powertrain Technologies; ERTRAC: Brussels, Belgium, 2016. 8. O’Connor, J.; Borz, M.; Ruth, D.; Han, J.; Paul, C.; Imren, A.; Haworth, D.; Martin, J.; Boehman, A.; Li, J.; et al. Optimization of an Advanced Combustion Strategy Towards 55% BTE for the Volvo SuperTruck Program. SAE Int. J. Engines 2017, 10, 1217–1227. [CrossRef] 9. Kocher, L. Final Scientific/Technical Report: Enabling Technologies for Heavy-Duty Vehicles—Cummins 55BTE. Available online: https://www.osti.gov/servlets/purl/1474075 (accessed on 1 March 2020). [CrossRef] 10. Castiglione, T.; Morrone, P.; Falbo, L.; Perrone, D.; Bova, S. Application of a Model-Based Controller for Improving Internal Combustion Engines Fuel Economy. Energies 2020, 13, 1148. [CrossRef] 11. Shkolnik, N.; Shknolnik, S. Rotary High Efficiency Hybrid Cycle Engine; SAE Technical Paper 2008-01-2448; SAE International: Warrendale, PA, USA, 2008. 12. Spreitzer, J.; Zahradnik, F.; Geringer, B. Implementation of a Rotary Engine (Wankel Engine) in a CFD Simulation Tool with Special Emphasis on Combustion and Flow Phenomena; SAE Technical Paper 2015-01-0382; SAE International: Warrendale, PA, USA, 2015. 13. Kutlar, O.A.; Malkaz, F. Two-Stroke Wankel Type Rotary Engine: A New Approach for Higher Power Density. Energies 2019, 12, 4096. [CrossRef] Energies 2020, 13, 2362 20 of 21

14. Li, Z.; Shih, T.I.P.; Schock, H.J.; Willis, E.A. A Numerical Study on the Effects of Apex Seal Leakage on Wankel; SAE Technical Papers; SAE International: Warrendale, PA, USA, 1991. 15. Gasim, M.M.; Noor, M.M.; Kadirgama, K. Theoretical Torque Comparison between Revetec Engine and Conventional Internal Combustion Engine. In Proceedings of the 1st National Conference in Mechanical Engineering Research and Postgraduate Students (NCMER 2010), Universiti Malaysia Pahang, Kuantan, Malaysia, 26–27 May 2010. 16. Meldolesi, R.; Badain, N. Scuderi Split Cycle Engine: Air Hybrid Vehicle Powertrain Simulation Study; SAE Technical Paper 2012-01-1013; SAE International: Warrendale, PA, USA, 2012. 17. Ashley, S. New Axial-Piston Engine Aimed at Range Extenders, Military Applications. Available online: https://www.sae.org/news/2014/08/new-axial-piston-engine-aimed-at-range-extenders-military- applications (accessed on 1 April 2020). 18. Zima, S.; Ficht, R. Ungewöhnliche Motoren (in German); Vogel Business Media: Würzburg, Germany, 2010. 19. Zhang, Y.; Zuo, Z.; Liu, J. Numerical Analysis on Combustion Characteristic of Leaf Spring Rotary Engine. Energies 2015, 8, 8086. [CrossRef] 20. Savvakis, S.; Gkoutzamanis, V.; Samaras, Z. Description of a Novel Concentric Rotary Engine; SAE Technical Paper 2018-01-0365; SAE International: Warrendale, PA, USA, 2018. [CrossRef] 21. Ferguson, C.R.; Kirkpatrick, A.T. Internal Combustion Engines—Applied Thermosciences; John Wiley & Sons: Hoboken, NJ, USA, 2001. 22. Blog, T. 8 September 2008. Available online: http://pressroom.toyota.com/article_print.cfm?article_id=2722 (accessed on 1 April 2020). 23. Heywood, J.B. Internal Combustion Engine Fundamentals; McGraw-Hill: New York, NY, USA, 1988. 24. Kobayashi, H.; Sugihara, H.; Yoshida, H.; Kawase, N.; Akimoto, Y. Effect of Top Rings on Piston Slap Noise; SAE Technical Paper 952236; SAE International: Warrendale, PA, USA, 1995. [CrossRef] 25. Gkoutzamanis, V.G.; Savvakis, S.; Kalfas, A.I. Numerical Simulation of a Novel Rotary Engine Compared to Conventional Reciprocating Engine Cycle. In Proceedings of the GPPS Conference 2017, Shanghai, China, 30 October–1 November 2017. 26. Zhang, M.; Yang, J.; Xu, W.; Xia, Y. Leakage and rotordynamic performance of a mixed labyrinth seal compared with that of a staggered labyrinth seal. J. Mech. Sci. Technol. 2017, 31, 2261–2277. [CrossRef] 27. Kang, Y.; Kim, T.S.; Kang, S.Y.; Moon, H.K. Aerodynamic Performance of Stepped Labyrinth Seals for Gas Turbine Applications. In Proceedings of the ASME Turbo Expo 2010: Power for Land, Sea, and Air. Volume 4: Heat Transfer, Parts A and B, Glasgow, UK, 14–18 June 2010; pp. 1191–1199. 28. Memmott, E.A. Empirical Estimation of a Load Related Cross-Coupled Stiffness and the Lateral Stability of Centrifugal Compressors. In Proceedings of the 18th Machinery Dynamics Seminar, Canadian Machinery Vibration Association, Halifax, NS, Canada, 26–28 April 2000. 29. Gao, R.; Kirk, G. CFD Study on Stepped and Drum Balance Labyrinth Seal. Tribol. Trans. 2013, 56, 663–671. [CrossRef] 30. Kim, T.S.; Cha, K.S. Comparative analysis of the influence of labyrinth seal configuration on leakage behavior. J. Mech. Sci. Technol. 2009, 23, 2830–2838. [CrossRef] 31. Eser, D.; Dereli, Y. Comparisons of Rotordynamic Coefficients in Stepped Labyrinth Seals by Using Colebrook-White Friction Factor Model. Meccanica 2007, 42, 177–186. [CrossRef] 32. Stoker, H.L. Determining and Improving Labyrinth Seal Performance in Current and Advanced High Performance Gas Turbines; AGARD–CP–237 (AGARD–AR–123), Paper 13; AGARD: London, UK, August 1978. 33. Waschka, W.; Wittig, S.; Kim, S. Influence of high rotational speeds on the heat transfer and discharge coefficients in labyrinth seals. J. Turbomach. 1992, 114, 462–468. [CrossRef] 34. Ha, T. Rotordynamic Analysis for Stepped-Labyrinth Gas Seals Using Moody’s Friction-Factor Model. J. Mech. Sci. Technol. 2001, 15, 1217–1225. [CrossRef] 35. Yuecel, U. Calculation of Leakage and Dynamic Coefficients of Stepped Labyrinth Gas Seals. Appl. Math. Comput. 2004, 152, 521–533. [CrossRef] 36. Subramaniana, S.; Sekhara, A. Rotordynamic characterization of rotating labyrinth gas turbine seals with radial growth_Combined centrifugal and thermal effects. Int. J. Mech. Sci. 2017, 123, 1–19. [CrossRef] 37. Mirzamoghadam, A.V.; Xiao, Z. Flow and heat transfer in an industrial rotor-stator rim sealing cavity. J. Eng. Gas Turbines Power 2002, 124, 125–132. [CrossRef] Energies 2020, 13, 2362 21 of 21

38. Thorat, M.; Childs, D. Predicted rotordynamic behavior of a labyrinth seal as rotor surface velocity approaches mach1. ASME J. Eng. Gas Turbines Power 2010, 132, 112504. [CrossRef] 39. Li, Z.; Li, J.; Feng, Z. Numerical comparisons of rotordynamic characteristics for three types of labyrinth gas seals with inlet preswirl. IMechE Part A J. Power Energy 2016.[CrossRef] 40. Converge Science, Converge CFD 2.4 Manual; CONVERGE CFD: Detroit, MI, USA, 2018. 41. Savvakis, S. Computational Investigation of a Novel Concentric Rotary Engine. Ph.D. Thesis, Aristotle University of Thessaloniki, Thessaloniki, Greece, 2011. 42. Schramm, V.; Denecke, J.; Kim, S.; Wittig, S. Shape Optimization of a Labyrinth Seal Applying the Simulated Annealing Method. Int. J. Rotating Mach. 2004, 10, 365–371. [CrossRef]

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